Vehicle control device and vehicle control method

ABSTRACT

A driving state estimator unit configured to detect a state quantity indicating vehicle body orientation, and a control unit configured to control vehicle body orientation using drive force from an engine when the absolute value of the amplitude of the detected state quantity is less than a second predetermined value, and using force generated by a second orientation control device instead of the drive force from the engine when the absolute value of the amplitude is equal to or greater than the second predetermined value.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a U.S. National stage application of InternationalApplication No. PCT/JP2013/051171, filed Jan. 22, 2013, which claimspriority to Japanese Patent Application No. 2012-012593 filed in Japanon Jan. 25, 2012, Japanese Patent Application No. 2012-012594 filed inJapan on Jan. 25, 2012, and Japanese Patent Application No. 2012-012595filed in Japan on Jan. 25, 2012, the contents of each of which arehereby incorporated herein by reference.

BACKGROUND

1. Field of the Invention

The present invention relates to a control device and a control methodfor controlling the state of a vehicle.

2. Background Information

Japanese Laid-Open Patent Application No. 2009-247157 discloses atechnique of controlling vehicle body orientation using drive force froma motive power source.

SUMMARY

However, a problem exists that controlling vehicle body orientationusing drive force from a motive power source when a state quantityindicating vehicle body orientation has a high amplitude will lead tolarge variations in vehicle drive force, creating an unnatural feel forthe driver.

An object of the present invention is to provide a vehicle controldevice and vehicle control method capable of reducing unnatural feelexperienced by a driver.

In order to achieve the abovementioned object, a vehicle control deviceaccording to the present invention controls vehicle body orientationusing a motive power source orientation control device when the absolutevalue of the amplitude of a detected state quantity is less than apredetermined value, and using a second orientation control deviceinstead of the motive power source orientation control device when theabsolute value of the amplitude is equal to or greater than thepredetermined value.

As a result, vehicle body orientation is not controlled using driveforce from a motive power source when amplitude is high, allowingvariations in drive force to be minimized and the level of unnaturalfeel experienced by a driver to be reduced.

BRIEF DESCRIPTIONS OF THE DRAWINGS

Referring now to the attached drawings which form a part of thisoriginal disclosure.

FIG. 1 is a schematic system diagram of a vehicle control deviceaccording to a first embodiment.

FIG. 2 is a control block diagram showing a configuration of controlperformed by the vehicle control device according to the firstembodiment.

FIG. 3 is a control block diagram showing a configuration of roll rateminimization control according to the first embodiment.

FIG. 4 is a time chart showing an envelope waveform formation processperformed in the roll rate minimization control of the first embodiment.

FIG. 5 is a control block diagram showing the configuration of a drivingstate estimator unit of the first embodiment.

FIG. 6 is a control block diagram showing the specifics of control in astroke speed calculator unit of the first embodiment.

FIG. 7 is a block diagram showing the configuration of a reference wheelspeed calculator unit of the first embodiment.

FIGS. 8A and 8B are schematic diagrams of a vehicle body vibrationmodel.

FIG. 9 is a control block diagram of actuator control amount calculationprocesses performed during pitch control in the first embodiment.

FIG. 10 is a control block diagram of brake pitch control in the firstembodiment.

FIG. 11 is a graph simultaneously showing a wheel speed frequencyprofile detected by a wheel speed sensor and a stroke frequency profilefrom a stroke sensor not installed in the present embodiment.

FIG. 12 is a control block diagram showing frequency-sensitive controlin sprung mass vibration damping control in the first embodiment.

FIG. 13 is a correlation graph showing human sensation profiles indifferent frequency regions.

FIG. 14 is a plot showing the relationship between the proportion ofvibration contamination and damping force in a float region in thefrequency-sensitive control of the first embodiment.

FIG. 15 is a wheel speed frequency profile detected by a wheel speedsensor in certain driving conditions.

FIG. 16 is a block diagram showing a control configuration for unsprungmass vibration damping control in the first embodiment.

FIG. 17 is a control block diagram showing a control configuration for adamping force control unit of the first embodiment.

FIG. 18 is a flow chart of a damping coefficient reconciliation processperformed during a standard mode in the first embodiment.

FIG. 19 is a flow chart of a damping coefficient reconciliation processperformed during a sports mode in the first embodiment.

FIG. 20 is a flow chart of a damping coefficient reconciliation processperformed during a comfort mode in the first embodiment.

FIG. 21 is a flow chart of a damping coefficient reconciliation processperformed during a highway mode in the first embodiment.

FIG. 22 is a time chart showing changes in damping coefficient whendriving on a hilly road surface and a bumpy road surface.

FIG. 23 is a flow chart of a driving state-based mode selection processperformed by a damping coefficient-reconciling unit of the firstembodiment.

FIG. 24 is a control block diagram of actuator control amountcalculation processes performed during pitch control in a secondembodiment.

DETAILED DESCRIPTION OF THE EMBODIMENTS First Embodiment

FIG. 1 is a schematic system diagram of a vehicle control deviceaccording to a first embodiment. A vehicle comprises an engine 1constituting a power source, brakes 20 for generating braking torque byapplying frictional force to the wheels (brakes corresponding toindividual wheels will be referred to hereafter as follows: front rightbrake: 20FR; front left brake: 20FL; rear right brake: 20RR; rear leftbrake: 20RL), and variable-damping-force shock absorbers 3 providedbetween each of the wheels and the vehicle body (“shock absorber” willbe abbreviated “S/A” in the following description; shock absorberscorresponding to individual wheels will be referred to as follows: frontright S/A: 3FR; front left S/A: 3FL; rear right S/A: 3RR; rear left S/A:3RL).

The engine 1 comprises an engine controller 1 a (also referred tohereafter as an engine control unit) for controlling the torqueoutputted by the engine 1; the engine controller 1 a controls the engineoperation state (engine rpm, engine output torque, etc.) as desired bycontrolling the opening of the throttle valve, the fuel injection level,the ignition timing, and the like of the engine 1. The brakes 20generates braking torque on the basis of hydraulic pressure suppliedfrom a brake control unit 2 capable of controlling brake hydraulicpressure for each of the wheels according to driving state. The brakecontrol unit 2 comprises a brake controller 2 a (also referred tohereafter as a brake control unit) for controlling the braking torquegenerated by the brakes 20; the desired hydraulic pressure is generatedin the brakes 20 for each of the wheels by the opening and closing of aplurality of solenoid valves using master cylinder pressure generated bya driver operating the brake pedal or pump pressure generated by abuilt-in motor-driven pump as a hydraulic pressure source.

The S/A 3 are damping force-generating devices for damping the elasticmotion of coil springs provided between the unsprung mass (the axles,wheels, etc.) and the sprung mass (vehicle body, etc.) of the vehicle,and the damping force generated thereby can be adjusted by the operationof actuators. Each of the S/As 3 comprises a cylinder in which fluid issealed, a piston that makes strokes within the cylinder, and an orificefor controlling the movement of the fluid between fluid chambers formedabove and below the piston. Orifices of various diameters are formed inthe piston, and an orifice corresponding to a control command isselected from the various orifices when the S/A actuator operates.Damping force corresponding to the diameter of the orifice is therebygenerated. The movement of the piston will be more easily restricted ifthe orifice diameter is small, increasing damping force, and movement ofthe piston will be less easily restricted if the orifice diameter islarge, decreasing damping force.

Apart from selecting the diameter of the orifice, damping force may alsobe set, for example, by disposing a solenoid control valve over apassage linking the fluid chambers formed above and below the piston andcontrolling the opening and closing of the solenoid control valve; theinvention is not particularly limited with respect thereto. Each of theS/As 3 comprises an S/A controller 3 a for controlling the damping forceof the S/A 3, and damping force is controlled by the operated of theorifice diameter by the S/A actuator.

Also comprised are wheel speed sensors 5 for detecting the wheel speedof each of the wheels (the sensors will be referred to as follows whenwheel speeds corresponding to individual wheels are indicated: frontright wheel speed: 5FR; front left wheel speed 5FL; rear right wheelspeed: SRR; rear left wheel speed: 5RL), an integrated sensor 6 fordetecting forward/reverse acceleration, yaw rate, and lateralacceleration acting upon the center of gravity of the vehicle, asteering angle sensor 7 for detecting a steering angle indicating theamount to which the driver has operated the steering wheel, a vehiclespeed sensor 8 for detecting vehicle speed, an engine torque sensor 9for detecting engine torque, an engine rpm sensor 10 for detectingengine rpm, a master pressure sensor 11 for detecting master cylinderpressure, a brake switch 12 for outputting an on state signal when abrake pedal is operated, and an accelerator opening sensor 13 fordetecting the degree to which an accelerator pedal is open. Signals fromthe various sensors are inputted to the S/A controller 3 a. Theintegrated sensor 6 may be disposed at the center of gravity of thevehicle or at another location without restriction as long as the sensoris capable of estimating various values at the position of the center ofgravity. The sensor need not be integrated; individual sensors fordetecting yaw rate, forward/reverse acceleration, and lateralacceleration may also be provided.

FIG. 2 is a control block diagram showing a configuration of controlperformed by the vehicle control device according to the firstembodiment. The first embodiment comprises three controllers: the enginecontroller 1 a, the brake controller 2 a, and the S/A controller 3 a.The S/A controller 3 a comprises a driver input control unit 31 forperforming driver input control so that a desired vehicle orientation isattained on the basis of driver operations (of the steering wheel,accelerator and brake pedals, etc.), a driving state estimator unit 32for estimating driving state on the basis of values detected by thesensors, a sprung mass vibration damping control unit 33 for controllingthe vibrational state of the sprung mass of the vehicle on the basis ofthe estimated driving state, an unsprung mass vibration damping controlunit 34 for controlling the vibrational state of the unsprung mass ofthe vehicle on the basis of the estimated driving state, and a dampingforce control unit 35 for deciding upon the damping force to be set forthe S/As 3 on the basis of a shock absorber orientation control amountoutputted from the driver input control unit 31, a sprung mass vibrationdamping control amount outputted from the sprung mass vibration dampingcontrol unit 33, and an unsprung mass vibration damping control amountoutputted from the unsprung mass vibration damping control unit 34, andcontrolling S/A damping force.

In the first embodiment, three controllers are provided, but the presentinvention is not particularly limited to such a configuration; forexample, the damping force control unit 35 may be provided separatelyfrom the S/A controller 3 a as an orientation controller for a total offour S/A controllers including the damping force control unit 35, or allthe controllers may be combined into a single integrated controller. Theconfiguration of the first embodiment envisions the repurposing of anengine controller and a brake controller in an existing vehicle as anengine control unit 1 a and a brake control unit 2 a, and theinstallation of a separate S/A controller 3 a to create the vehiclecontrol device according to the first embodiment.

(Overall Configuration of Vehicle Control Device)

Three actuators are used in the vehicle control device according to thefirst embodiment in order to control the vibrational state of the sprungmass of the vehicle. Because the control performed by each of theactivators affects the state of the sprung mass of the vehicle,interference is a problem. In addition, the elements controllable by theengine 1, the elements controllable by the brakes 20, and the elementscontrollable by the S/As 3 all differ, and the matter of thecombinations in which these elements should be controlled is anotherproblem. For example, the brakes 20 are capable of controlling bouncingmotion and pitching motion, but controlling both at the same time willcreate a strong sense of deceleration and tend to create an unnaturalfeel for the driver. The S/As 3 are capable of absorbing rolling motion,bouncing motion, and pitching motion, but controlling all three of theseusing the S/As 3 will lead to increased manufacturing costs for the S/As3, and tends to increase damping force; this facilitates high-frequencyvibrational input from the road surface, also creating an unnatural feelfor the driver. In other words, a trade-off must be made in that controlperformed by the brakes 20 will not lead to worse high-frequencyvibration but will lead to an increased sense of deceleration, andcontrol performed by the S/As 3 will not create a sense of decelerationbut will lead to high-frequency vibrational input.

Thus, a control configuration has been adopted for the vehicle controldevice of the first embodiment in which a comprehensive assessment ismade of these problems in order to draw upon the respective advantagesof these control methods so as to complement the weaknesses of theother, thereby yielding a vehicle control device that is both economicaland offers superior vibration damping performance. To this end, thefollowing points were taken into consideration in the overallconstruction of the control system.

(1) The operation of the actuators with regard to pitch control isselectively turned on and off according to the amplitude of the statequantity indicating vehicle body orientation (pitch rate in the firstembodiment), thereby striking a better balance in terms of theabovementioned trade-off.

(2) The only type of motion subjected to control by the brakes 20 ispitching motion, thereby eliminating the sense of deceleration producedfrom control by the brakes 20.

(3) The amount of control performed by the engine 1 and the brakes 20 isrestricted to less than the actually outputtable control amount, therebyreducing the burden placed upon the S/As 3 and minimizing the unnaturalfeel yielded by control performed by the engine 1 and the brakes 20.

(4) Skyhook control is performed by all of the actuators. This allowsskyhook control to be inexpensively performed using all of the wheelspeed sensors installed in the vehicle, without the use of a strokesensor, sprung mass vertical acceleration sensor, or the like, as isusually necessary to perform skyhook control.

(5) Scalar control (frequency-sensitive control) has been newlyintroduced in order to address high-frequency vibrational input, whichis difficult to address using skyhook control or other types of vectorcontrol, when the S/As 3 are performing sprung mass control.

(6) The control state manifested by the S/As 3 is selected, asappropriate, according to the driving state, thereby providing a controlstate suited to the driving conditions.

The foregoing has been a summary of the features of the control systemaccording to the embodiment as a whole. The specifics by which each ofthese individual features will be described in sequence hereafter.

(Driver Input Control Unit)

First, the driver input control unit will be described. The driver inputcontrol unit 31 comprises an engine driver input control unit 31 a forattaining the vehicle orientation demanded by the driver by controllingthe torque of the engine 1, and an S/A driver input control unit 31 bfor attaining the vehicle orientation demanded by the driver bycontrolling the damping force of the S/As 3. The engine driver inputcontrol unit 31 a calculates a ground load variation minimizationcontrol amount for minimizing variations in the ground loads of thefront wheels and the rear wheels, and a yaw response control amountcorresponding to the vehicle behavior desired by the driver on the basisof signals from the steering angle sensor 7 and the vehicle speed sensor8.

The S/A driver input control unit 31 b calculates a driver input dampingforce control amount corresponding to the vehicle behavior desired bythe driver on the basis of the signals from the steering angle sensor 7and the vehicle speed sensor 8, and outputs this amount to the dampingforce control unit 35. If, for example, the nose of the vehicle riseswhile the driver is turning, the driver's field of view can easily betaken off the road; in such cases, the damping force of the four wheelsis outputted as the driver input damping force control amount so as toprevent the nose from rising. A driver input damping force controlamount for minimizing rolling generating during turning is alsooutputted.

(Controlling Rolling Via S/A Driver Input Control)

Roll minimization control performed via S/A driver input control willnow be described. FIG. 3 is a control block diagram showing aconfiguration of roll rate minimization control according to the firstembodiment. A lateral acceleration estimator unit 31 b 1 estimateslateral acceleration Yg on the basis of a front wheel steering angle δfdetected by the steering angle sensor 7, a rear wheel steering angle δr(the actual rear wheel steering angle if a rear wheel steering device isprovided; otherwise, 0 as appropriate), and vehicle speed VSP detectedby the vehicle speed sensor 8. The lateral acceleration Yg is calculatedaccording to the following formula using an estimated yaw rate value γ.

Yg=VSp·γ

The estimated yaw rate value γ is calculated according to the followingformula:

$\begin{Bmatrix}\beta \\\gamma\end{Bmatrix} = {N\begin{Bmatrix}\delta_{f} \\\delta_{r}\end{Bmatrix}}$ $\begin{Bmatrix}\beta \\\gamma\end{Bmatrix} = {M^{- 1}N\begin{Bmatrix}\delta_{f} \\\delta_{r}\end{Bmatrix}}$ wherein ${M = \begin{bmatrix}m_{11} & m_{12} \\m_{21} & m_{22}\end{bmatrix}},{N = \begin{bmatrix}n_{11} & n_{12} \\n_{21} & n_{22}\end{bmatrix}}$ m₁₁ = −(Ktf ⋅ Lf − Ktv ⋅ Lv)$m_{12} = {{- \frac{1}{V}}\left( {{{Ktf} \cdot {Lf}^{2}} - {{Ktv} \cdot {Lv}^{2}}} \right)}$m₂₁ = −2(Ktf + Ktv)$m_{22} = {{{- \frac{2}{V}}\left( {{{Ktf} \cdot {Lf}} - {{Ktv} \cdot {Lv}}} \right)} - {M \cdot V}}$n₁₁ = −Ktf ⋅ Lf n₁₂ = Ktv ⋅ Lr n₂₁ = −2 ⋅ Ktf n₂₂ = −2 ⋅ Ktv

vehicle body slide angle: β

vehicle body yaw rate: γ

front wheel steering angle: δf

rear wheel steering angle: δr

vehicle body: V

front wheel CP: Ktf

rear wheel CP: Ktv

distance between front axle and center of gravity: Lf

distance between rear axle and center of gravity: Lr

vehicle body mass: M

A 90° phase lead component-generating unit 31 b 2 differentiates theestimated lateral acceleration Yg and outputs a lateral accelerationderivative dYg. A 90° phase lag component-generating unit 31 b 3 outputsa component F(dYg) in which the phase of the lateral accelerationderivative dYg is delayed 90°. In the component F(dYg) 90°, the phase ofthe component from which the low-frequency region has been removedgenerated by the phase lead component-generated unit 31 b 2 is returnedto the phase of the lateral acceleration Yg, and the DC is cut from thelateral acceleration Yg; i.e., the component is a transitional componentof lateral acceleration Yg. A 90° phase lag component-generating unit 31b 4 outputs a component F(Yg) in which the phase of the lateralacceleration Yg is delayed 90°. A gain multiplier unit 31 b 5 multipliesthe lateral acceleration Yg, lateral acceleration derivative dYg,lateral acceleration DC-cut component F(dYg), and 90° phase lagcomponent F(Yg) by gain. Gain is set according to the roll rate transferfunction for the steering angle. Gain may also be adjusted according tofour control modes described hereafter. A square calculator unit 31 b 6squares and outputs the components having been multiplied by gain. Asynthesizer unit 31 b 7 adds the values outputted by the squaringprocessor unit 31 b 6. A gain multiplier unit 31 b 8 multiplies thesquare of the summed components by gain, and outputs the results. Asquare root calculator unit 31 b 9 calculates the square root of thevalue outputted by the gain multiplier unit 31 b 7, thereby calculatinga driver input orientation control amount for use in roll rateminimization control, and outputs the amount to the damping forcecontrol unit 35.

The 90° phase lead component-generating unit 31 b 2, 90° phase lagcomponent-generating unit 31 b 3, 90° phase lag component-generatingunit 31 b 4, gain multiplier unit 31 b 5, square calculator unit 31 b 6,synthesizer unit 31 b 7, gain multiplier unit 31 b 8, and square rootcalculator unit 31 b 9 constitute a Hilbert transform unit 31 b 10 forgenerating an envelope waveform using a Hilbert transform.

FIG. 4 is a time chart showing an envelope waveform formation processperformed in the roll rate minimization control of the first embodiment.When a driver begins steering at time t1, a roll rate gradually beginsto be generated. At this point, the 90° phase lead component dYg isadded to form an envelope waveform, and the driver input orientationcontrol amount is calculated on the basis of the envelope waveform-basedscalar quantity, thereby allowing roll rate generation during theinitial stage of steering to be minimized. In addition, by adding thelateral acceleration DC-cut component F(dYg) to form the envelopewaveform, roll rates generated during transitional states, such as whenthe driver is beginning or ending steering, can be efficientlyminimized. In other words, damping force is not excessively increasedand degradation in ride comfort can be avoided during constant turningstates in which roll is stably generated.

Next, when the driver holds the steering wheel in place at time t2, the90° phase lead component dYG and the lateral acceleration DC-cutcomponent F(dYg) disappear, and the 90° phase lag component F(Yg) isthen added. At this time, a roll rate resonance component equivalent tothe roll aftershock is generated after rolling occurs, even if there islittle change in the roll rate itself in a steady steering state. If thephase lag component F(Yg) were not added, damping force would be set toa low value from time t2 to time t3, risking destabilization of vehiclebehavior by the roll rate resonance component. The 90° phase lagcomponent F(Yg) is added in order to minimize the roll rate resonancecomponent.

When the driver shifts from the held steering state back to astraight-ahead driving state at time t3, the lateral acceleration Ygdecreases, and the roll rate is reduced to a low value. The action ofthe 90° phase lag component F(Yg) also ensures damping force at thispoint as well, allowing destabilization due to the roll rate resonancecomponent to be avoided.

(Driving State Estimator Unit)

Next, the driving state estimator unit will be described. FIG. 5 is acontrol block diagram showing the configuration of a driving stateestimator unit of the first embodiment. The driving state estimator unit32 of the first embodiment calculates a stroke speed, bounce rate, rollrate, and pitch rate for each wheel used in the skyhook controlperformed by the sprung mass vibration damping control unit 33 asdescribed hereafter primarily on the basis of the wheel speeds detectedby the wheel speed sensors 5. The values from the wheel speed sensors 5of the wheels are inputted into a stroke speed calculator unit 321, andsprung mass speed is calculated by the stroke speed calculator unit 321from the stroke speeds calculated for the wheels.

FIG. 6 is a control block diagram showing the specifics of control in astroke speed calculator unit of the first embodiment. A stroke speedcalculator unit 321 is separately provided for each wheel; the controlblock diagram shown in FIG. 6 focuses on a specific wheel. The strokespeed calculator unit 321 comprises a reference wheel speed calculatorunit 300 for calculating a reference wheel speed on the basis of thevalues from the wheel speed sensors 5, the front wheel steering angle δfdetected by the steering angle sensor 7, a rear wheel steering angle δr(the actual rear wheel steering angle if a rear wheel steering device isprovided, 0 otherwise), a vehicle body lateral speed, and an actual yawrate detected by the integrated sensor 6, a tire rotational vibrationfrequency calculator unit 321 a for calculating tire rotationalvibration frequency on the basis of the calculated reference wheelspeed, a deviation calculator unit 321 b for calculating the deviationbetween the reference wheel speed and the value from the wheel speedsensor (i.e., wheel speed variation), a GEO conversion unit 321 c forconverting the deviation calculated by the deviation calculator unit 321b to a suspension stroke amount, a stroke speed calibrator unit 321 dfor calibrating the converted stroke amount to a stroke speed, and asignal processing unit 321 e for applying a band elimination filtercorresponding to the frequency calculated by the tire rotationalvibration frequency calculator unit 321 a to the calibrated valueyielded by the stroke speed calibrator unit 321 d to eliminate a primarytire rotational vibration component and calculate a final stroke speed.

(Reference Wheel Speed Calculator Unit)

The reference wheel speed calculator unit 300 will now be described.FIG. 7 is a block diagram showing the configuration of a reference wheelspeed calculator unit of the first embodiment. The reference wheel speedis a wheel speed from which various types of interference from theindividual wheels have been removed. In other words, the differencebetween the value from the wheel speed sensor and the reference wheelspeed is related to a component that varies according to a strokegenerated by vehicle body bouncing motion, rolling motion, pitchingmotion, or unsprung vertical vibration; in the present embodiment, thestroke speed is calculated on the basis of this difference.

A flat surface motion component extractor unit 301 uses the wheel speedsensor values as inputs to calculate a first wheel speed V0 as areference wheel speed for each of the wheels on the basis of the vehiclebody plan view model. ω (rad/s) is the wheel speed sensor detected bythe wheel speed sensor 5, δf (rad) is a front wheel actual steeringangle detected by the steering angle sensor 7, δr (rad) is a rear wheelactual steering angle, Vx is vehicle body lateral speed, γ (rad/s) isthe yaw rate detected by the integrated sensor 6, V (m/s) is a vehiclebody speed estimated from the calculated reference wheel speed ω0, VFL,VFR, VRL, and VRR are the reference wheel speeds to be calculated, Tf isa front wheel treat, Tr is a rear wheel treat, Lf is the distance fromthe position of the vehicle center of gravity to the front wheels, andLr is the distance from the position of the vehicle center of gravity tothe rear wheel. The vehicle body plan view model is expressed as followsusing the symbols described above.

VFL=(V−Tf/2·γ)cos δf+(Vx+Lf·γ)sin δf

VFR=(V+Tf/2·γ)cos δf+(Vx+Lf·γ)sin δf

VRL=(V−Tr/2·γ)cos δr+(Vx−Lr·γ)sin δr

VRR=(V+Tr/2·γ)cos δr+(Vx−Lr·γ)sin δr  (Formula 1)

If a normal driving state in which no lateral sliding of the vehicleoccurs is hypothesized, 0 may be inputted for the vehicle body lateralspeed Vx. This yields the following formulas when the various formulasare rewritten with values based on V. When rewriting in this manner, Vis written as V0FL, V0FR, V0RL, and V0RR (equivalent to first wheelspeeds) as values corresponding to the various wheels.

V0FL={VFL—L·γ sin δf}/cos δf+Tf/2·γ

V0FR={VFR−Lf·γ sin δf}/cos δf−Tf/2·γ

V0RL={VRL+Lr·γ sin δr}/cos δr+Tr/2·γ

V0RR={VRR+Lf·γ sin δf}/cos δR−Tr/2·γ  (Formula 2)

A roll interference-removing unit 302 uses the first wheel speed V0 asan input to calculate second wheel speeds V0F, V0R as reference wheelspeeds for the front and rear wheels on the basis of a vehicle bodyfront view model. The vehicle body front view model is used to removewheel speed differences generated by rolling motion occurring around acenter of roll rotation on a normal line passing through the vehiclecenter of gravity as viewed from the front of the vehicle, and isrepresented as follows.

V0F=(V0FL+V0FR)/2

V0R=(V0RL+V0RR)/2

This yields second wheel speeds V0F, V0R from which roll-basedinterference has been removed.

A pitch interference-removing unit 303 uses the second wheel speeds V0F,V0R as inputs to calculate third wheel speeds VbFL, VbFR, VbRL, and VbRRconstituting reference wheel speeds for all the wheels according to avehicle body side view model. The vehicle body side view model is usedto remove wheel speed differences generated by pitching motion occurringaround a center of pitch rotation on a normal line passing through thevehicle center of gravity as viewed from the side of the vehicle.

VbFL=VbFR=VbRL=VbRR={Lr/(Lf+Lr)}V0F+{Lf/(Lf+Lr)}V0R  (Formula 3)

A reference wheel speed redistribution unit 304 uses VbFL(=VbFR=VbRL=VbRR) for V in the vehicle body plan view model shown informula 1 to calculate final reference wheel speeds VFL, VFR, VRL, VRRfor each wheel, which are divided by the tire radius r0 to calculate thereference wheel speed ω0.

Once the reference wheel speed ω0 has been calculated according to theprocess described above, the deviation between the reference wheel speedω0 and the wheel speed sensor value is calculated; the deviationrepresents a wheel speed variation arising from suspension strokes, andis therefore converted into a stroke speed Vzs. As a rule, not only doesa suspension make strokes in the vertical direction when holding thewheels, but the wheel rotational centers move forwards and backwards asstrokes occur, and the axles equipped with the wheel speed sensors 5become tilted, creating a difference in rotational angle with thewheels. Because this forward and backward motion leads to changes inwheel speed, deviations between the reference wheel speed and the wheelspeed sensor value can be extracted as stroke-induced variations. Thedegree of variation that occurs can be set, as appropriate, according tothe suspension geometry.

Once the stroke speed calculator unit 321 has calculated the strokespeeds Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR for each wheel according to theprocess described above, a sprung mass speed calculator unit 322calculates a bounce rate, roll rate, and pitch rate for use in skyhookcontrol.

(Estimation Model)

In skyhook control, damping force is set according to the relationshipbetween the stroke speeds of the S/As 3 and the sprung mass speed, andthe orientation of the sprung mass is controlled to achieve a flatdriving state. In order to achieve control of the orientation of thesprung mass via skyhook control, feedback on the sprung mass speed isnecessary. Stroke speed is a value detectable from the wheel speedsensor 5; since a sensor for the vertical acceleration of the sprungmass is not provided, the sprung mass speed must be estimated using anestimation model. Problems involved in the estimation model and theappropriate model configuration to adopt will now be discussed.

FIGS. 8A and 8B are schematic diagrams of a vehicle body vibrationmodel. FIG. 8A is a model for a vehicle provided with S/As of constantdamping force (hereafter referred to as a conventional vehicle), andFIG. 8B is a model for a vehicle provided with variable S/As in whichskyhook control is performed. In FIGS. 8A and 8B, Ms indicates sprungmass, Mu indicates unsprung mass, Ks indicates coil spring modulus ofelasticity, Cs indicates S/A damping coefficient, Ku indicates unsprung(tire) modulus of elasticity, Cu indicates unsprung (tire) dampingcoefficient, and Cv indicates a variable damping coefficient. z2indicates the position of the sprung mass, z1 indicates the position ofthe unsprung mass, and z0 indicates the position of the road surface.

If the conventional vehicle model shown in FIG. 8A is used, the equationof motion for the sprung mass is expressed as follows. The first-orderdifferential for z1 (i.e., speed) is represented by dz1, and thesecond-order differential (i.e., acceleration) is represented by ddz1.

Ms·ddz2=−Ks(z2−z1)−Cs(dz2−dz1)  (Estimation formula 1)

Applying a Laplace transform to this relationship yields the followingformula.

dz2=−(1/Ms)·(1/S ²)·(Cs·s+Ks)(dz2−dz1)  (Estimation formula 2)

Because dz2−dz1 is stroke speed (Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR), thesprung mass speed can be calculated from the stroke speed. However,adjusting damping force via skyhook control will vastly reduceestimation precision, creating the problem that a large orientationcontrol force (damping force adjustment) cannot be applied in theconventional vehicle model.

Thus, the use of a skyhook control-based vehicle model shown in FIG. 8Bis conceivable. As a rule, altering damping force involves altering theforce limiting the piston movement speed of the S/As 3 as suspensionstrokes occur. Because semi-active S/As 3 in which the pistons cannot beactively moved in a desired direction, a semi-active skyhook model isused; calculating sprung mass speed yields the following formula.

dz2=−(1/Ms)·(1/s ²)·{(CS+CV)·s+Ks}(dz2−dz1)  (Estimation formula 3)

wherein:

if dz2·(dz2−dz1)>0, Cv=Csky·{dz21(dz2−dz1)}, and

if dz2·(dz2−dz1)<0, Cv=0.

That is, Cv is a discontinuous value.

If the semi-active skyhook model is viewed as a filter when one wishesto estimate sprung mass speed using a simple filter, the variables areequivalent to filter coefficients, and a variable damping coefficient Cvthat is discontinuous in the pseudo-differential term {(Cs+Cv)·s+Ks};thus, filter response is unstable, and suitable estimation precisioncannot be obtained. In particular, unstable filter response will lead tophase shifting. Skyhook control cannot be achieved if the correspondencebetween the phase and the sign for sprung mass speed breaks down. Thedecision was thus made to estimate sprung mass speed using an activeskyhook model, in which a stable Csky can be directly used withoutrelying upon the sign relationship between sprung mass speed and thestroke speed even if semi-active S/As 3 are used. The use of an activeskyhook model to calculate sprung mass speed can be expressed asfollows.

dz2=−(1/s)·{1/(s+Csky/Ms)}·{(Cs/Ms)s+(Ks/Ms)}(dz2−dz1)  (Estimationformula 4)

In this case, there is no discontinuity in the pseudo-differential term{(Cs/Ms)s+(Ks/Ms)}, and the term {1/(s+Csky/Ms)} can be constitutedusing a low-pass filter. Filter response is therefore stable, andsuitable estimation precision is obtainable. It should be noted that,even if an active skyhook model is adopted, only semi-active control isactually possible; thus, the controllable range is halved. The estimatedsprung mass speed is therefore less than the actual speed in thefrequency band from sprung mass resonance down. However, it is the phasethat is most vital in the context of skyhook control, and skyhookcontrol can be achieved as long as the correspondence between phase andsign can be maintained; this is unproblematic because the sprung massspeed can be adjusted using the other coefficients or the like.

It is apparent from the relationship described above that sprung massspeed can be estimated if the stroke speeds for each wheel are known.Because an actual vehicle has not one wheel, but four, we will nowconsider using the stroke speeds for each wheel to estimate the state ofthe sprung mass divided into roll rate, pitch rate, and bounce ratemodes. When calculating the abovementioned three components from thestroke speeds of the four wheels, one corresponding component islacking, leading to an indefinite solution; thus, warp rate, whichindicates the movement of diagonally opposed wheels, has beenintroduced. Defining xsB as the bounce term of the stroke amount, xsR asthe roll term, xsP as the pitch term, xsW as the warp term, and z_sFL,z_sFR, z_sRL, z_sRR as stroke amounts corresponding to Vz_sFL, Vz_sFR,Vz_sRL, Vz_sRR, the following formula arises.

$\begin{matrix}\begin{matrix}{\begin{Bmatrix}{z\_ sFL} \\{z\_ sFR} \\{z\_ sRL} \\{z\_ sRR}\end{Bmatrix} = {\begin{bmatrix}1 & 1 & {- 1} & {- 1} \\1 & {- 1} & {- 1} & 1 \\1 & 1 & 1 & 1 \\1 & {- 1} & 1 & {- 1}\end{bmatrix}\left\{ \begin{matrix}{xsB} \\{xsR} \\{xsP} \\{xsW}\end{matrix} \right\rbrack \begin{Bmatrix}{xsB} \\{xsR} \\{xsP} \\{xsW}\end{Bmatrix}}} \\{= {\begin{bmatrix}1 & 1 & {- 1} & {- 1} \\1 & {- 1} & {- 1} & 1 \\1 & 1 & 1 & 1 \\1 & {- 1} & 1 & {- 1}\end{bmatrix}^{- 1}\begin{Bmatrix}{z\_ sFL} \\{z\_ sFR} \\{z\_ sRL} \\{z\_ sRR}\end{Bmatrix}}}\end{matrix} & \left( {{Formula}\mspace{14mu} 1} \right)\end{matrix}$

In view of the relationship shown above, the differentials dxsB, . . .of xsB, xsR, xsP, xsW are expressed by the following formulas.

dxsB=1/4(Vz _(—) sFL+Vz _(—) sFR+Vz _(—) sRL+Vz _(—) sRR)

dxsR=1/4(Vz _(—) sFL−Vz _(—) sFR+Vz _(—) sRL−Vz _(—) sRR)

dxsP=1/4(−Vz _(—) sFL−Vz _(—) sFR+Vz _(—) sRL+Vz _(—) sRR)

dxsW=1/4(−Vz _(—) sFL+Vz _(—) sFR+Vz _(—) sRL−Vz _(—) sRR)

The relationship between sprung mass speed and stroke speed is obtainedfrom estimation formula 4 above; thus, taking G as the section ofestimation formula 4 reading—(1/s)·{1/(s+Csky/Ms)}·{(Cs/Ms)s+(Ks/Ms)},GB, GR, and GP as values taking into account modal parameters (CskyB,CskyR, CskyP, CsB, CsR, CsP, KsB, KsR, KsP) for the bounce terms, rollterms, and pitch terms of Csky, Cs, and Ks, respectively, dB as bouncerate, dR as roll rate, and dP as pitch rate, dB, dR, and dP can becalculated as follows.

dB=GB·dxsB

dR=GR·dxsR

dP=GP·dxsP

As shown from the foregoing, the state of the sprung mass of an actualvehicle can be estimated on the basis of the stroke speeds for thevarious wheels.

(Sprung Mass Vibration Damping Control Unit)

Next, the configuration of the sprung mass vibration damping controlunit 33 will be described. As shown in FIG. 2, the sprung mass vibrationdamping control unit 33 comprises a skyhook control unit 33 a forcontrolling orientation according to the estimated value for sprung massspeed described above, and a frequency-sensitive control unit 33 b forminimizing sprung mass vibration on the basis of the road surface inputfrequency.

(Configuration of Skyhook Control Unit)

The vehicle control device according to the first embodiment comprisesthree actuators for achieving sprung mass orientation control in theform of the engine 1, the brakes 20, and the S/As 3. Of these, bouncerate, roll rate, and pitch rate are the objects of control for the S/As3, bounce rate and pitch rate are the objects of control for the engine1, and pitch rate is the object of control for the brakes 20 in theskyhook control unit 33 a. In order to allocate control amounts to aplurality of actuators that act in different manners and control thestate of the sprung mass, a shared control amount must be used for each.In the first embodiment, the control amount for each of the actuatorscan be determined by using the sprung mass speed estimated by thedriving state estimator unit 32.

Bounce-directional skyhook control amount:

FB=CskyB·dB

Roll-directional skyhook control amount:

FR=CskyR·dR

Pitch-directional skyhook control amount:

FP=CskyP·dP

FB is sent to the engine 1 and the S/As 3 as the bounce orientationcontrol amount. FR is for control performed only by the S/As 3, and sois sent to the damping force control unit 35 as a roll orientationcontrol amount.

Next, the pitch-directional skyhook control amount FP will be described.Pitch control is performed by the engine 1, the brakes 20, and the S/As3. FIG. 9 is a control block diagram of actuator control amountcalculation processes performed during pitch control in the firstembodiment. The skyhook control unit 33 a comprises a first targetorientation control amount calculator unit 331 for calculating a targetpitch rate constituting a first target orientation control amount thatis a control amount that can used in common for all of the actuators, anengine orientation control amount calculator unit 332 for calculatingthe engine orientation control amount achieved by the engine 1, a brakeorientation control amount calculator unit 334 for calculating the brakeorientation control amount achieved by the brakes 20, an S/A orientationcontrol amount calculator unit 336 for calculating the S/A orientationcontrol amount achieved by the S/As 3, and an operation switching unit337 for selectively turning the operation of the actuators with respectto pitch control on and off.

As the skyhook control according to the present system gives foremostpriority to pitch rate minimization, the first target orientationcontrol amount calculator unit 331 outputs pitch rate without furthermodification (hereafter, this pitch rate will be referred to as thefirst target orientation control amount).

The pitch rate calculated by the driving state estimator unit 32 isinputted into the operation switching unit 337. If the absolute value ofthe amplitude of the pitch rate is less than a first predeterminedvalue, the first target orientation control amount outputted from thefirst target orientation control amount calculator unit 331 is outputtedto the engine orientation control amount calculator unit 332. If theabsolute value of the amplitude of the pitch rate is less than a secondpredetermined value that is greater than the first predetermined value,the first target orientation control amount is outputted to the S/Aorientation control amount calculator unit 336. If the absolute value ofthe amplitude of the pitch rate is equal to or greater than the secondpredetermined value, the first target orientation control amount isoutputted to the brake orientation control amount calculator unit 334.

The engine orientation control amount 332 calculator unit calculates theengine orientation control amount, which is the control amountachievable by the engine 1, on the basis of the inputted first targetorientation control amount. A limit value that limits the engine torquecontrol amount according to the engine orientation control amount so asnot to create an unnatural feel for the driver is set in the engineorientation control amount calculator unit 332. The engine torquecontrol amount is thus kept within a predetermined forward/reverseacceleration range when converted to forward/reverse acceleration. Theengine control unit 1 a calculates the engine torque control amount onthe basis of the engine orientation control amount corresponding to thelimit value, and outputs this amount to the engine 1.

The brake orientation control amount calculator unit 334 calculates thebrake orientation control amount, which is the control amount achievableby the brakes 20, on the basis of the inputted first target orientationcontrol amount. A limit value for limiting the braking torque controlamount so as not to create an unnatural feel for the driver, as in thecase of the engine 1, is set in the brake orientation control amountcalculator unit 334 (this limit value will be discussed in detailhereafter). The braking torque control amount is thus kept within apredetermined forward/reverse acceleration range when converted toforward/reverse acceleration (with the limit value being calculated onthe basis of naturalness of ride feel for passengers, actuator lifespan,etc.). The brake control unit 2 a calculates a braking torque controlamount (or deceleration) on the basis of the brake orientation controlamount corresponding to the limit value, and is outputted to the brakecontrol unit 2.

The S/A orientation control amount calculator unit 336 calculates thepitch orientation control amount for the S/As 3 on the basis of theinputted first target orientation control amount. The damping forcecontrol unit 35 calculates the damping force control amount on the basisof the bounce orientation control amount, the roll orientation controlamount, and the pitch orientation control amount (hereafter collectivelyreferred to as the S/A orientation control amounts), and outputs thisamount to the S/As 3.

(Brake Pitch Control)

Brake pitch control will now be described. Generally, the brakes 20 arecapable of controlling both bounce and pitch; thus, it is preferablethat they control both. However, when bounce control is performed by thebrakes 20, braking force is applied to all four wheels simultaneously,and there is a strong sense of deceleration even in directions of lowcontrol priority despite the difficulty in obtaining control effects,tending to create an unnatural feed for the driver. Thus, aconfiguration in which the brakes 20 specialize in pitch control hasbeen adopted. FIG. 10 is a control block diagram of brake pitch controlin the first embodiment. Defining m as the mass of the vehicle body, Bffas front wheel braking force, BFr as rear wheel braking force, Hcg asthe height between the vehicle center of gravity and the road surface, aas vehicle acceleration, Mp as pitch moment, and Vp as pitch rate, thefollowing relationships hold.

BFf+BFr=m·a

m·a·Hcg=Mp

Mp=(BFf+BFr)·Hcg

If braking force is applied when the pitch rate Vp is positive, i.e.,the front wheel side of the vehicle is lowered, the front wheel sidewill sink further lower, augmenting pitch motion; thus, braking force isnot applied in such cases. On the other hand, when the pitch rate Vp isnegative, i.e., the front wheel side of the vehicle is raised, thebraking pitch moment will impart braking force, minimizing the rising ofthe front wheel side. This ensures the driver's field of view and makesthe area ahead easier to see, contributing to improved senses of safetyand flatness of ride. In other words, the following control amounts areapplied:

when Vp>0 (front wheels lowered), Mp=0; and

when Vp≦0 (front wheels raised), Mp=CskyP·Vp

Braking torque is thus generated only when the front side of the vehicleis raised, thereby allowing the sense of deceleration created thereby tobe reduced compared to cases in which braking torque is generated bothwhen the front side is raised and when it is lowered. In addition, theactuators need only be operated at half the frequency as usual, allowinginexpensive actuators to be used.

The brake orientation control amount calculator unit 334 comprises thefollowing control blocks on the basis of the relationship describedabove. A dead band process sign determiner unit 3341 determines the signfor the inputted pitch rate Vp; if the sign is positive, no control isnecessary, so 0 is outputted to a deceleration sense-reducing processor3342, and if the sign is negative, control is determined to be possible,so a pitch rate signal is outputted to the deceleration sense-reducingprocessor 3342.

(Deceleration Sense Reduction Process)

Next, a deceleration sense reduction process will be described. Thisprocess corresponds to the limit applied by the limit value set in thebrake orientation control amount calculator unit 334. A squaringprocessor 3342 a squares the pitch rate signal. This reverses the sign,and smoothes the increase in control force. A pitch rate square dampingmoment calculator unit 3342 b multiplies the squared pitch rate by askyhook gain CskyP for the pitch term that takes the squaring processinto account to calculate pitch moment Mp. A target decelerationcalculator unit 3342 c divides the pitch moment Mp by mass m and theHeight Hcg between the vehicle center of gravity and the road surface tocalculate a target deceleration.

A jerk threshold value limiter unit 3342 d determines whether the rateof change in the calculated target deceleration, i.e., jerk, is withinpreset deceleration jerk threshold value and release jerk thresholdvalue ranges, and whether the target deceleration is within aforward/reverse acceleration limit value range. If any of the thresholdvalues is exceeds, the target deceleration is corrected to a valuewithin the ranges for the jerk threshold values. If the targetdeceleration exceeds the limit value, it is set to within the limitvalue. It is thereby possible to generate deceleration so as not tocreate an unnatural feel for the driver.

A target pitch moment converter unit 3343 multiplies the targetdeceleration limited by the jerk threshold value limiter unit 3342 d bythe mass m and the height Hc9 to calculate a target pitch moment, whichis outputted to the brake control unit 2 a and a target pitch rateconverter unit 3344. The target pitch rate converter unit 3344 dividesthe target pitch moment by the pitch term skyhook gain CskyP to convertto a target pitch rate (equivalent to a brake orientation controlamount), which is outputted to a third target orientation control amountcalculator unit 335.

In the first embodiment, the action of the operation switching unit 337causes pitch control to be performed by the engine 1 alone when theabsolute value of the amplitude of the pitch rate is less than the firstpredetermined value, by the S/As 3 only when the absolute value of theamplitude of the pitch rate is equal to or greater than the firstpredetermined value and less than the second predetermined value, and bythe brakes 20 alone when the absolute value of the amplitude of thepitch rate is equal to or greater than the second predetermined value.

In other words, pitch control is not performed by the S/As 3 when thepitch rate is low (i.e., the absolute value of the amplitude of thepitch rate is less than the first predetermined value), thereby allowingthe controllable range of the S/As 3 to be reduced, and enabling pitchcontrol to be achieved using inexpensive S/As 3. As a rule, an increasein damping force control amount at this point would lead to increaseddamping force. Because increased damping force leads to stiff suspensioncharacteristics, high-frequency input is more easily transmitted whenthere is high-frequency vibration input from the road surface, reducingpassenger comfort (this situation is hereafter referred to asexacerbated high-frequency vibration characteristics). By contrast, notperforming pitch control using the S/As 3 allows exacerbation ofhigh-frequency vibration to be avoided when the pitch rate is low.

Pitch control is not performed by the brakes 20 when the pitch rate ismiddling (i.e., the absolute value of the amplitude of the pitch rate isequal to or greater than the first predetermined value and less than thesecond predetermined value) or low, thereby allowing an increased senseof deceleration arising from increased braking torque to be avoided whenpitch rate is middling or low. Because pitch rate is rarely high (i.e.,the absolute value of the amplitude of the pitch rate is equal to orgreater than the second predetermined value), the number of situationsin which deceleration is generated can be reduced, improving thedurability of the brake system.

In addition, not performing pitch control using the S/As 3 when thepitch rate is high allows the controllable range of the S/As 3 to bereduced, allowing for pitch control using inexpensive S/As 3.Exacerbated high-frequency vibration in situations in which the pitchrate is high can also be avoided.

(Frequency-Sensitive Control Unit)

Next, a frequency-sensitive control process performed in the sprung massvibration damping control unit will be described. In the firstembodiment, as a rule, the sprung mass speed is estimated on the basisof the values detected by the wheel speed sensors 5, and skyhook controlis performed based thereon, thereby achieving sprung mass vibrationdamping control. However, there are cases in which it may not bepossible to guarantee sufficient estimation precision using the wheelspeed sensors 5, and cases in which it is desirable to activelyguarantee a comfortable driving state (i.e., a soft ride rather than afeeling of vehicle body flatness) depending on driving conditions or thedriver's intent. In such cases, it may be difficult to effect suitablecontrol due to slight phase shifts if vector control such as skyhookcontrol, in which the relationship (phase, etc.) of the signs of thestroke speed and the sprung mass speed is vital, is used; thus,frequency-sensitive control constituted by sprung mass vibration dampingcontrol according to vibration profile scalar quantities has beenintroduced.

FIG. 11 is a graph simultaneously showing a wheel speed frequencyprofile detected by a wheel speed sensor and a stroke frequency profilefrom a stroke sensor not installed in the present embodiment. In thiscontext, “frequency profile” refers to a profile in which the magnitudeof amplitude against frequency is plotted on the y axis. A comparison ofthe frequency component of the wheel speed sensor 5 and the frequencycomponent of the stroke sensor shows that roughly similar scalarquantities can be plotted from the sprung mass resonance frequencycomponent to the unsprung mass resonance frequency component. Thus, thedamping force has been set on the basis of this frequency profile out ofthe values detected by the wheel speed sensor 5. The region in which thesprung mass resonance frequency component is present is a frequencyregion in which the swaying of a passenger's entire body creates asensation as thought the passenger is floating in the air, that is, thatthe gravitational acceleration affecting the passenger has decreased,and is referred to as the float region (0.5-3 Hz). The region betweenthe sprung mass resonance frequency component and the unsprung massresonance frequency component is a frequency region in which, althoughthere is no sensation of reduced gravitational acceleration, there is asensation similar to the quick, frequent bouncing experienced by aperson on horseback when riding at a trot, that is, an up-and-downmotion followed by the entire body, and is referred to as the bounceregion (3-6 Hz). The region in which the unsprung mass resonancefrequency component is present is a frequency region in which, althoughvertical body mass movement is not experienced, frequent vibrations areconveyed to a part a passenger's body, such as the thighs, and isreferred to as a flutter region (6-23 Hz).

FIG. 12 is a control block diagram showing frequency-sensitive controlin sprung mass vibration damping control in the first embodiment. A bandelimination filter 350 cuts out noise apart from the frequency componentin the wheel speed sensor value that is used to perform control. Apredetermined frequency region splitter unit 351 splits the frequencycomponent into a float region, a bounce region, and a flutter region. AHilbert transform processor unit 352 performs a Hilbert transform uponthe split frequency bands, converting them to scalar quantities(specifically, areas calculated using amplitude and frequency band)based on the amplitude of the frequency. A vehicle vibrational systemweight-setting unit 353 sets weights for the vibration from the floatregion, bounce region, and flutter region frequency bands that isactually propagated to the vehicle. A human sensation weight-settingunit 354 sets weights for the vibration from the float region, bounceregion, and flutter region frequency bands that is actually propagatedto a passenger.

The setting of human sensation weights will now be described. FIG. 13 isa correlation graph showing human sense profiles plotted againstfrequency. As shown in FIG. 13, passenger sensitivity to frequencies iscomparatively low in the low-frequency float region, with sensitivitygradually increasing as one moves into regions of higher frequencies.Frequencies in the flutter region and higher-frequency regions becomeprogressively harder to transmit to the passenger. In view of this, thefloat region human sensation weight Wf is set at 0.17, the bounce regionhuman sensation weight Wh is set higher than Wf at 0.34, and the flutterregion human sensation weight Wb is set higher than Wf and Wh at 0.38.It is thereby possible to increase the correlation between the scalarquantities for the various frequency bands and the vibration actuallypropagated to a passenger. These two weighting factors may be altered,as appropriate, according to vehicle concept or passenger preferences.

A weight-determining device 355 calculates the proportions occupied bythe weight for each of the frequency bands. Defining a as the floatregion weight, b as the bounce region weight, and c as the flutterregion weight, the weighting factor for the float region is (a/(a+b+c)),the weighting factor for the bounce region is (b/(a+b+c)), and theweighting factor for the flutter region is (c/(a+b+c)). A scalarquantity calculator unit 356 multiplies the scalar quantities for thevarious frequency bands calculated by the Hilbert transform processorunit 352 by the device 355, and outputs final scalar quantities. Theprocess up to this point is performed on the wheel speed sensor valuesfor each of the wheels.

A maximum value-selecting unit 357 selects the maximum value out of thefinal scalar quantities calculated for each of the four wheels. The 0.01at the bottom is set to avoid having 0 as a denominator, as the total ofthe maximum values is used as a denominator in a subsequent process. Aproportion calculator unit 358 calculates proportions using the total ofthe maximum scalar quantity values for each of the frequency bands asthe denominator and the maximum scalar quantity value for the frequencyband equivalent to a float region as the numerator. In other words, theproportion of contamination (hereafter, simply “proportion”) in thefloat region contained in all vibration components is calculated. Asprung mass resonance filter 359 performs a filter process having asprung mass resonance frequency of roughly 1.2 Hz on the calculatedproportion, and extracts a sprung mass resonance frequency bandcomponent representing the float region from the calculated proportion.In other words, because the float region is present at roughly 1.2 Hz,it is believed that the proportion of this region will also vary around1.2 Hz. The finally extracted proportion is then outputted to thedamping force control unit 35, and a frequency-sensitive damping forcecontrol amount corresponding to the proportion is outputted.

FIG. 14 is a plot showing the relationship between the proportion ofvibration contamination and damping force in a float region in thefrequency-sensitive control of the first embodiment. As shown in FIG.14, a high damping force is set when the float region occupies a largeproportion, thereby reducing the vibration level of sprung massresonance. Even if a high damping force is set, no high-frequencyvibration or bounce-like vibration is transmitted to the passengerbecause the bounce region and flutter region occupy small proportions.Meanwhile, setting a low damping force when the float region occupies asmall proportion reduces the vibration transmission profile at and abovesprung mass resonance, minimizing high-frequency vibration and yieldinga smooth ride.

The benefits of frequency-sensitive control in a comparison offrequency-sensitive control and skyhook control will now be described.FIG. 15 is a wheel speed frequency profile detected by a wheel speedsensor 5 in certain driving conditions. This profile especially appearswhen driving on road surfaces having continuous small bumps, such ascobbled roads. When skyhook control is performed while driving on a roadsurface exhibiting this profile, the problem arises that, becausedamping force is determined using the peak amplitude value in skyhookcontrol, any degradation in phase estimation for high-frequencyvibrational input will cause an extremely high damping force to be setat the wrong timing, leading to exacerbated high-frequency vibration.

By contrast, if control is performed using scalar quantities rather thanvectors, as in frequency-sensitive control, the float region occupies asmall proportion on road surfaces such as that shown in FIG. 15, leadingto a low damping force being set. Thus, even if the amplitude of flutterregion vibration is high, the vibration transmission profile issufficiently reduced, allowing the exacerbation of high-frequencyvibration to be avoided. As shown by the foregoing, high-frequencyvibration can be minimized via scalar quantity-based frequency-sensitivecontrol in regions where control is difficult due to degradations inphase estimation precision even if skyhook control is performed using anexpensive sensor.

(Unsprung Mass Vibration Damping Control Unit)

Next, the configuration of the unsprung mass vibration damping controlunit will be described. As discussed in the context of the conventionalvehicle shown in FIG. 8B, a resonance frequency band is also present intires, as they possess both a modulus of elasticity and a dampingcoefficient. However, because a tire has a mass that is less than thatof the sprung mass, and a high modulus of elasticity as well, the bandis present toward frequencies higher than sprung mass resonance. Theunsprung mass resonance component causes tire rumbling in the unsprungmass, potentially degrading ground contact. In addition, rumbling in theunsprung mass can be uncomfortable for passengers. Thus, damping forceis set according to the unsprung mass resonance component in order tominimize unsprung mass resonance-induced rumbling.

FIG. 16 is a block diagram showing a control configuration for unsprungmass vibration damping control in the first embodiment. An unsprung massresonance component extractor unit 341 applies a band-pass filter to thewheel speed variation outputted from the deviation calculator unit 321 bof the driving state estimator unit 32 to extract an unsprung massresonance component. The unsprung mass resonance component is extractedfrom the region at roughly 10-20 Hz in the wheel speed frequencycomponent. An envelope waveform-forming unit 342 scalarizes theextracted unsprung mass resonance component, and forms an envelopewaveform using an envelope filter. A gain multiplier unit 343 multipliesthe scalarized unsprung mass resonance component by the gain, calculatesan unsprung mass vibration damping force control amount, which isoutputted to the damping force control unit 35. In the first embodiment,an unsprung mass resonance component is extracted by applying aband-pass filter to the wheel speed variation outputted from thedeviation calculator unit 321 b of the driving state estimator unit 32,but it is also acceptable to apply a band-pass filter to the valuedetected by the wheel speed sensor to extract the unsprung massresonance component, or for the driving state estimator unit 32 toestimate the unsprung mass speed along with the sprung mass speed toextract an unsprung mass resonance component.

(Configuration of Damping Force Control Unit)

Next, the configuration of the damping force control unit 35 will bedescribed. FIG. 17 is a control block diagram showing a controlconfiguration for a damping force control unit of the first embodiment.The driver input damping force control amount outputted from the driverinput control unit 31, the S/A orientation control amount outputted fromthe skyhook control unit 33 a, the frequency-sensitive damping forcecontrol amount outputted from the frequency-sensitive control unit 33 b,the unsprung mass vibration damping force control amount outputted fromthe unsprung mass vibration damping control unit 34, and the strokespeed calculated by the driving state estimator unit 32 are inputtedinto an equivalent viscous damping coefficient converter unit 35 a,which converts these values into an equivalent viscous dampingcoefficient.

A damping coefficient-reconciling unit 35 b reconciles which dampingcoefficient, out of the damping coefficients converted by the equivalentviscous damping coefficient converter unit 35 a (hereafter referred toindividually as the driver input damping coefficient k1, the S/Aorientation damping coefficient k2, the frequency-sensitive dampingcoefficient k3, and the unsprung mass vibration damping coefficient k4),is used to perform control, and outputs a final damping coefficient. Acontrol signal converter unit 35 c converts a control signal (commandedcurrent value) to be sent to the S/As 3 on the basis of the dampingcoefficient reconciled by the damping coefficient-reconciling unit 35 band the stroke speed, and outputs the signal to the S/As 3.

(Damping Coefficient-Reconciling Unit)

Next, the specifics of the reconciliation performed by the dampingcoefficient-reconciling unit 35 b will be described. The vehicle controldevice of the first embodiment has four control modes. The first mode isstandard mode, for situations in which suitable steering conditions areobtainable while driving on general urban roads. The second mode issports mode, for situations in which stable steering conditions areavailable while aggressively driving along winding roads and the like.The third mode is comfort mode, for situations in which priority isgiven to comfort while driving, such as when starting off at low vehiclespeeds. The fourth mode is highway mode, for situations involvingdriving at high vehicle speeds on highways and the like with multiplestraight sections.

In standard mode, priority is given to unsprung mass vibration dampingcontrol performed by the unsprung mass vibration damping control unit 34while skyhook control is being performed by the skyhook control unit 33a. In sports mode, skyhook control is performed by the skyhook controlunit 33 a and unsprung mass vibration damping control is performed bythe unsprung mass vibration damping control unit 34 while givingpriority to driver input control performed by the driver input controlunit 31. In comfort mode, priority is given to unsprung mass vibrationdamping control performed by the unsprung mass vibration damping controlunit 34 while frequency-sensitive control is being performed by thefrequency-sensitive control unit 33 b. In highway mode, the controlamount for the unsprung mass vibration damping control performed by theunsprung mass vibration damping control unit 34 is added to the skyhookcontrol performed by the skyhook control unit 33 a while given priorityto the driver input control performed by the driver input control unit31. Damping coefficient reconciliation in these various modes will nowbe described.

(Reconciliation in Standard Mode)

FIG. 18 is a flow chart of a damping coefficient reconciliation processperformed during a standard mode in the first embodiment.

In step S1, it is determined whether the S/A orientation dampingcoefficient k2 is greater than the unsprung mass vibration dampingcoefficient k4, and, if so, the process continues to step S4, and k2 isset as the damping coefficient.

In step S2, the scalar quantity proportion of the flutter region iscalculated on the basis of the scalar quantities for the float region,bounce region, and flutter region described in the context of thefrequency-sensitive control unit 33 b.

In step S3, it is determined whether the proportion of the flutterregion is equal to or greater than a predetermined value, and, if so,the process continues to step S4 and the low value k2 is set as thedamping coefficient for fear of high-frequency vibration reducing ridecomfort. On the other hand, if the proportion of the flutter region isless than the predetermined value, there is no worry of high-frequencyvibration reducing ride comfort even if a high damping coefficient isset, so the process continues to step S5, and k4 is set as thecoefficient.

In standard mode, as discussed above, priority is given, as a rule, tounsprung mass vibration damping control, which minimizes resonance inthe unsprung mass. However, if the damping force required for skyhookcontrol is less than the damping force required for unsprung massvibration damping control, and the flutter region occupies a largeproportion, the damping force for skyhook control is set so as to avoidexacerbating the high-frequency vibration profile in order to meet therequirements of unsprung mass vibration damping control. This allows anoptimal damping profile to be obtained according to the driving state,allowing high-frequency vibration-induced degradations of ride comfortto be avoided while simultaneously achieving a flat vehicle body feel.

(Reconciliation in Sports Mode)

FIG. 19 is a flow chart of a damping coefficient reconciliation processperformed during a sports mode in the first embodiment.

In step S11, the damping force distribution factors for the four wheelsare calculated on the basis of the driver input damping coefficients k1for the four wheels set during driver input control. Defining k1fr asthe front right wheel driver input damping coefficient, k1fl as thefront left wheel driver input damping coefficient, k1rr as the rearright wheel driver input damping coefficient, k1rl as the rear leftwheel driver input damping coefficient, and xfr, xfl, xrr, and xrl asthe damping force distribution factors for the different wheels, thedistribution factors are calculated as follows:

xfr=k1fr/(k1fr+k1fl+k1rr+k1rl)

xfl=k1fl/(k1fr+k1fl+k1rr+k1rl)

xrr=k1rr(k1fr+k1fl+k1rr+k1rl)

xrl=k1rl/(k1fr+k1fl+k1rr+k1rl)

In step S12, it is determined whether a damping force distributionfactor x is within a predetermined range (greater than α and less thanβ), and, if so, distribution is determined to be roughly equal for allthe wheels, and the process continues to step S13; if even one factor isoutside the predetermined range, the process continues to step S16.

In step S13, it is determined whether the unsprung mass vibrationdamping coefficient k4 is greater than the driver input dampingcoefficient k1, and, if so, the process continues to step S15, and k4 isset as a first damping coefficient k. On the other hand, if the unsprungmass vibration damping coefficient k4 is equal to or less than thedriver input damping coefficient k1, the process continues to step S14,and k1 is set as the first damping coefficient k.

In step S16, it is determined whether the unsprung mass vibrationdamping coefficient k4 is the maximum value max that can be set for theS/As 3; if so, the process continues to step S17, and, if not, theprocess continues to step S18.

In step S17, the maximum value for the driver input damping coefficientsk1 for the four wheels is the unsprung mass vibration dampingcoefficient k4, and the damping coefficient that satisfies the dampingforce distribution factor is calculated as the first damping coefficientk. In other words, a value is calculated such that the dampingcoefficient is maximized while the damping force distribution factor issatisfied.

In step S18, a damping coefficient such that the damping forcedistribution factor is satisfied within a range in which the driverinput damping coefficients k1 for all four wheels are equal to orgreater than k4. In other words, a value is calculated such that thedamping force distribution factor set by the driver input control issatisfied, and the requirements of unsprung mass vibration dampingcontrol are also met.

In step S19, it is determined whether the first damping coefficients kset in the abovementioned steps are less than the S/A orientationdamping coefficient k2 set during skyhook control; if so, k2 is set andthe process continues to step S20 due to the damping coefficientrequired by skyhook control being larger. On the other hand, if k isequal to or greater than k2, k is set and the process continues to stepS21.

In sports mode, as discussed above, priority is given, as a rule, tounsprung mass vibration damping control, which minimizes resonance inthe unsprung mass. However, because the damping force distributionfactor required by driver input control is closely related to thevehicle body orientation, and is particularly deeply related to changesin driver line of view caused by roll mode, foremost priority is givento ensuring the damping force distribution factor, rather than thedamping coefficient required by driver input control itself. Formovement that causes changes in vehicle body orientation whilepreserving the damping force distribution factor, selecting skyhookcontrol via select high allows a stable vehicle body orientation to bemaintained.

(Reconciliation in Comfort Mode)

FIG. 20 is a flow chart of a damping coefficient reconciliation processperformed during a comfort mode in the first embodiment.

In step S30, it is determined whether the frequency-sensitive dampingcoefficient k3 is greater than the unsprung mass vibration dampingcoefficient k4, and, if so, the process continues to step S32 and thefrequency-sensitive damping coefficient k3 is set. On the other hand, ifthe frequency-sensitive damping coefficient k3 is determined to be equalto or less than the unsprung mass vibration damping coefficient k4, theprocess continues to step S32 and the unsprung mass vibration dampingcoefficient k4 is set.

In comfort mode, as discussed above, priority is given, as a rule, tounsprung mass resonance damping control, which minimizes resonance inthe unsprung mass. Because frequency-sensitive control is performed assprung mass vibration damping control to begin with, and an optimaldamping coefficient for the road surface conditions is set, control canbe performed while ensuring ride comfort, allowing sensationsinsufficient ground contact caused by rattling in the unsprung mass tobe avoided through unsprung mass vibration damping control. In comfortmode, as in standard mode, it is acceptable for the damping coefficientto be switched according to the proportion of flutter in the frequencyscalar quantity. This allows for a super comfort mode in which ridecomfort is even better ensured.

(Reconciliation in Highway Mode)

FIG. 21 is a flow chart of a damping coefficient reconciliation processperformed during a highway mode in the first embodiment. The samereconciliation process as in sports mode is performed from steps S11 toS18; thus, description thereof will be omitted. In step S40, the S/Aorientation damping coefficient k2 yielded by skyhook control is addedto the reconciled first damping coefficient k yielded by the process upto step S18 and outputted.

In highway mode, as discussed above, the sum of the reconciled firstdamping coefficient k and the S/A orientation damping coefficient k2 isused to reconcile the damping coefficient. This operation will now bedescribed with reference to the drawings. FIG. 22 is a time chartshowing changes in damping coefficient when driving on a hilly roadsurface and a bumpy road surface. For instance, if an attempt is made tominimize swaying motion in the vehicle body caused by the effects ofslight hills in the road surface when driving at high vehicle speeds viaskyhook control alone, it is necessary to detect slight variations inwheel speed, which requires that a comparatively high skyhook controlgain be set. In such cases, swaying motion can be minimized, but bumpsin the road surface can lead to the control gain being too great,creating the risk of excessive damping force control being performed.This gives rise to concerns of degraded ride comfort or vehicle bodyorientation.

By contrast, because the first damping coefficient k is constantly set,as in highway mode, a certain level of damping force can be constantlyensured, allowing swaying motion in the vehicle body to be minimizedeven if a low damping coefficient is used in skyhook control. Inaddition, because there is no need to increase the skyhook control gain,bumps in the road surface can be managed using a normal control gain.Moreover, because skyhook control is performed in a state in which thedamping coefficient k is set, a process of reducing the dampingcoefficient can be operated in a semi-active control region, unlike inthe case of a damping coefficient limit, ensuring a stable vehicleorientation during high-speed driving.

(Mode Selection Process)

Next, a mode selection process for selecting among the various drivingmodes described above will be described. FIG. 23 is a flow chart of adriving state-based mode selection process performed by a dampingcoefficient-reconciling unit of the first embodiment.

In step S50, it is determined whether the vehicle is driving straightahead based on the value from the steering angle sensor 7; if so, theprocess continues to step S51, and if the vehicle is determined to be ina state of turning, the process continues to step S54.

In step S51, it is determined whether the vehicle speed is equal to orgreater than a predetermined vehicle speed VSP1 indicating a state ofhigh vehicle speed on the basis of the value from the vehicle speedsensor 8, and, if so, the process continues to step S52 and standardmode is selected. On the other hand, if the speed is less than VSP1, theprocess continues to step S53 and comfort mode is selected.

In step S54, it is determined whether the vehicle speed is equal to orgreater than a predetermined vehicle speed VSP1 indicating a state ofhigh vehicle speed on the basis of the value from the vehicle speedsensor 8, and, if so, the process continues to step S55 and highway modeis selected. On the other hand, if the speed is less than VSP1, theprocess continues to step S56 and sports mode is selected.

That is, standard mode is selected when driving at a high vehicle speedwhen driving straight ahead, thereby making it possible to stabilize thevehicle body orientation via skyhook control, ensure ride comfort byminimizing high-frequency vibration-induced bouncing or fluttering, andminimizing resonance in the unsprung mass. Selecting comfort mode whendriving at low speeds makes it possible to minimize resonance in theunsprung mass while minimizing the transmission of vibration such asbouncing or fluttering to passengers.

Meanwhile, highway mode is selected when driving at a high vehicle speedin a state of turning, thereby performing control using a value to whicha damping coefficient has been added; thus, high damping force isyielded as a rule. It is thus possible to minimize unsprung massresonance while actively ensuring the unsprung mass resonance duringturning via driver input control, even when traveling at a high vehiclespeed. Selecting sports mode when driving at a low vehicle speed allowsunsprung mass resonance to be minimized while actively ensuring thevehicle body orientation during turning via driver input control andperforming skyhook control as appropriate, thereby allowing for drivingwith a stable vehicle orientation.

In the first embodiment, an example of a mode selection process in Whichthe driving state is detected and the mode is automatically switched hasbeen presented, but it is also possible to provide a mode switch or thelike that can be operated by a driver to select the driving mode. Thisyields ride comfort and turning performance matching the driver'sdesired driving state.

As described above, the first embodiment yields the following effects.

(1) Provided are an engine orientation control amount calculator unit332 (a motive power source orientation control device) for calculatingan engine orientation control amount for an engine 1 of a vehicleserving as a first orientation control device for obtaining a targetorientation for vehicle body orientation and outputting the amount tothe engine 1, a second orientation control device for calculating asecond orientation control amount for a second orientation controldevice for obtaining a target orientation for the vehicle bodyorientation and outputting the amount to the second orientation controldevice, a driving state estimator unit 32 (a state quantity detectiondevice) for detecting a state quantity indicating vehicle bodyorientation, and a skyhook control unit 33 a (an orientation controldevice) for controlling vehicle body orientation using the engineorientation control amount calculator unit 332 when the absolute valueof the amplitude of a detected state quantity is less than a secondpredetermined value, and controlling vehicle body orientation using thesecond orientation control device instead of the engine orientationcontrol amount calculator unit 332 when the absolute value of theamplitude is equal to or greater than the second predetermined value.

As a result, vehicle body orientation is not controlled using driveforce from the engine 1 when amplitude is high, allowing variations indrive force to be minimized and the level of unnatural feel experiencedby a driver to be reduced.

(2) The second orientation control device is the brakes 20, and thesecond orientation control device is a brake orientation control amountcalculator unit 334 (a friction brake orientation control device) forcalculating the brake orientation control amount for the brakes 20 as asecond orientation control amount and outputting this amount to thebrakes 20.

It is thereby possible to control orientation using the engine 1 and thebrakes 20, which are actuators that are completely unrelated to theexacerbation of high-frequency vibration characteristics, allowing suchexacerbation to be avoided. In other words, it is possible to controlorientation using the engine 1 and the brakes 20, which are actuatorsthat are completely unrelated to the exacerbation of high-frequencyvibration characteristics, allowing such exacerbation to be avoided. Inaddition, because orientation is controlled using only the engineorientation control amount for the engine 1 and the brake orientationcontrol amount for the brakes 20 is set to zero when the absolute valueof the amplitude of the state quantity representing vehicle bodyorientation is less than the second predetermined value, the number ofsituations in which deceleration is generated during vehicle bodyorientation control can be reduced, improving the durability of thebrake system.

(3) Provided is a S/A orientation control amount calculator unit 336 (adamping force control device) for calculating a shock absorberorientation control amount for the S/As 3 serving as a third orientationcontrol device in order that will yield a target orientation for thevehicle body orientation and outputting the amount to the S/As 3. Theskyhook control unit 33 a (an orientation control device) controlsvehicle body orientation using the engine orientation control amountcalculator unit 332 when the absolute value of the amplitude of thedetected state quantity is less than a first predetermined value, usingthe S/A orientation control amount calculator unit 336 instead of theengine orientation control amount calculator unit 332 when the absolutevalue of the amplitude is equal to or greater than the firstpredetermined value and less than a second predetermined value that isgreater than the first predetermined value, and using the brakeorientation control amount calculator unit 334 instead of the S/Aorientation control amount calculator unit 336 when the absolute valueof the amplitude is equal to or greater than the second predeterminedvalue.

As a result, the shock absorber orientation control amount for the S/As3 is zero when the absolute value of the amplitude of the state quantityrepresenting vehicle body orientation is less than the firstpredetermined value or equal to or greater than the second predeterminedvalue, enabling the controllable range of the S/As 3 to be reduced, andallowing vehicle body orientation to be controlled using an inexpensivearrangement. In addition, pitch control is performed by the S/As 3 in anarrow amplitude region, allowing the exacerbation of high-frequencyvibration profiles to be avoided. In addition, vehicle body orientationis not performed by the brakes 20 when the absolute value of theamplitude of the state quantity indicating vehicle body orientation isless than the second predetermined value, allowing the number ofsituations in which deceleration is generated to be reduced, andimproving the durability of the brake system.

(4) The second orientation control device is the S/As 3, and the secondorientation control device is the S/A orientation control amountcalculator unit 336 (a damping force control device) for calculating ashock absorber orientation control amount for the S/As 3 as a secondorientation control amount, and outputting the amount to the S/As 3.

As a result, the shock absorber orientation control amount for the S/As3 is zero when the absolute value of the amplitude of the state quantityrepresenting vehicle body orientation is less than the firstpredetermined value, enabling the controllable range of the S/As 3 to bereduced, and allowing vehicle body orientation to be controlled using aninexpensive arrangement.

(5) Provided is a brake orientation control amount calculator unit 334(a friction brake orientation control device) for calculating a brakeorientation control amount for the brakes 20 serving as a thirdorientation control device in order that will yield a target orientationfor vehicle body orientation and outputting the amount to the brakes 20.The skyhook control unit 33 a (an orientation control device) controlsvehicle body orientation using the engine orientation control amountcalculator unit 332 when the absolute value of the amplitude of thedetected state quantity is less than a first predetermined value, usingthe S/A orientation control amount calculator unit 336 instead of theengine orientation control amount calculator unit 332 when the absolutevalue of the amplitude is equal to or greater than the firstpredetermined value and less than a second predetermined value that isgreater than the first predetermined value, and using the brakeorientation control amount calculator unit 334 instead of the S/Aorientation control amount calculator unit 336 when the absolute valueof the amplitude is equal to or greater than the second predeterminedvalue.

As a result, the shock absorber orientation control amount for the S/As3 is zero when the absolute value of the amplitude of the state quantityrepresenting vehicle body orientation is less than the firstpredetermined value or equal to or greater than the second predeterminedvalue, enabling the controllable range of the S/As 3 to be reduced, andallowing vehicle body orientation to be controlled using an inexpensivearrangement. In addition, pitch control is performed by the S/As 3 in anarrow amplitude region, allowing the exacerbation of high-frequencyvibration profiles to be avoided. In addition, vehicle body orientationis not performed by the brakes 20 when the absolute value of theamplitude of the state quantity indicating vehicle body orientation isless than the second predetermined value, allowing the number ofsituations in which deceleration is generated to be reduced, andimproving the durability of the brake system.

(6) A driving state estimator unit 32 (a driving state detection device)for detecting the pitch rate of the vehicle is comprised, and the brakeorientation control amount calculator unit 334 calculates the brakeorientation control amount on the basis of the detected pitch rate.

Generally, the brakes 20 are capable of controlling both bounce andpitch; thus, it is preferable that they control both. However, whenbounce control is performed by the brakes 20, braking force is appliedto all four wheels simultaneously, and there is a strong sense ofdeceleration even in directions of low control priority despite thedifficulty in obtaining control effects, tending to create an unnaturalfeed for the driver. Thus, a configuration in which the brakes 20specialize in pitch control has been adopted.

If braking force is applied when the pitch rate Vp is positive, i.e.,the front wheel side of the vehicle is lowered, the front wheel sidewill sink further lower, augmenting pitch motion; thus, braking force isnot applied in such cases. On the other hand, when the pitch rate Vp isnegative, i.e., the front wheel side of the vehicle is raised, thebraking pitch moment will impart braking force, minimizing the rising ofthe front wheel side. This ensures the driver's field of view and makesthe area ahead easier to see, contributing to improved senses of safetyand flatness of ride. Because braking torque is generated only when thefront side of the vehicle body is raised, the amount of decelerationgenerated can be reduced compared to arrangements in which brakingtorque is generated both when the front side is raised and when it islowered. In addition, the actuators need only be operated at half thefrequency as usual, allowing inexpensive actuators to be used.

(7) The driving state estimator unit 32 (a driving state detectiondevice) estimates the pitch rate of the vehicle on the basis of changesin wheel speed.

The pitch rate is thus estimated using the wheel speed sensors 5 withwhich all vehicles are generally equipped, without the need forexpensive sensors such as a sprung weight vertical acceleration sensoror a stroke sensor, thereby allowing costs and the number of parts to bereduced, and the ease with which components are installed in the vehicleto be improved.

(8) The driving state estimator unit 32 (a driving state detectiondevice) estimates vehicle pitch rate using a four-wheel modelconstructed using a bounce term representing the four-wheel verticalmovement, a pitch term representing front and rear wheel verticalmovement, a roll term representing left and right wheel verticalmovement, and a warp term representing diagonal wheel vertical movement.

In other words, one corresponding component is lacking when the sprungmass speed for all four wheels is divided into roll term, pitch term,and bounce term modes, destabilizing the solution. Thus, the warp termrepresenting the movement of diagonal wheels is introduced in order toallow the abovementioned terms to be estimated.

(9) The engine orientation control amount calculator unit 332 (a motivepower source orientation control device), brake orientation controlamount calculator unit 334 (a friction brake orientation controldevice), and S/A orientation control amount calculator unit 336 (adamping force control device) calculate the various orientation controlamounts on the basis of skyhook control.

That is, applying skyhook control-based control amounts to the engine 1,brakes 20, and S/As 3 allows a stable sprung mass orientation to beobtained.

(10) The driving state estimator unit 32 (a state quantity detectiondevice) estimates vehicle pitch rate. As a result, the engineorientation control amount is set to zero when the pitch rate is high,allowing variations in drive force to be minimized, and reducing thelevel of unnatural feel experienced by the driver. In addition, thepitch orientation control amount of the S/As 3 is set to zero when thepitch rate is low, allowing the roll orientation control amount andbounce orientation control amount to be increased all the more, andimproving ease of skyhook control.

(11) Provided are a driving state estimator unit 32 (sensor) fordetecting a state quantity indicating vehicle body orientation, and askyhook control unit 33 a (controller) for controlling vehicle bodyorientation using drive force from the engine 1 when the absolute valueof the amplitude of the detected state quantity is less than the secondpredetermined value, and using force generated by the second orientationcontrol device instead of the drive force from the engine 1 when theabsolute value of the amplitude is equal to or greater than the secondpredetermined value.

As a result, vehicle body orientation is not controlled using driveforce from the engine 1 when amplitude is high, allowing variations indrive force to be minimized and the level of unnatural feel experiencedby a driver to be reduced.

(12) The skyhook control unit 33 a controls vehicle body orientationusing drive force from the engine 1 when the absolute value of theamplitude of the state quantity representing vehicle body orientation isless than the second predetermined value, and using force generated bythe second orientation control device instead of the engine 1 when theabsolute value of the amplitude is equal to or greater than the secondpredetermined value.

As a result, vehicle body orientation is not controlled using driveforce from the engine 1 when amplitude is high, allowing variations indrive force to be minimized and the level of unnatural feel experiencedby a driver to be reduced.

Second Embodiment

FIG. 24 is a control block diagram of actuator control amountcalculation processes performed during pitch control in a secondembodiment. The second embodiment differs from the first embodiment inthat the operation switching unit 337 switches the operation of theactuators on and off on the basis of the roll rate instead of the pitchrate.

In the second embodiment, the action of the operation switching unit 337causes pitch control to be performed using the engine torque controlamount alone when the absolute value of the amplitude of the roll rateis less than a first predetermined value, using the damping forcecontrol amount instead of the engine torque control amount when theabsolute value of the amplitude of the roll rate is equal to or greaterthan the first predetermined value and less than the secondpredetermined value, and using the braking torque control amount insteadof the damping force control amount when the absolute value of theamplitude of the roll rate is equal to or greater than the secondpredetermined value.

Thus, the second embodiment yields the following effect in addition tothe effects (1)-(9), (11), and (12) yielded by the first embodiment.

(13) The driving state estimator unit 32 (a state quantity detectiondevice) is a device for detecting the roll rate of the vehicle.

As a result, the engine orientation control amount is set to zero whenthe roll rate is high, allowing variations in drive force to beminimized, and reducing the level of unnatural feel experienced by thedriver. In addition, the brake orientation control amount is minimizedwhen the roll rate is high or low, allowing the roll orientation controlamount of the S/As 3 to be increased all the more, and allowing rollingmotion to be minimized at an early stage.

Other Embodiments

The foregoing has been a description of embodiments of the presentinvention with reference to the drawings, but the specific configurationof the present invention is not limited to these embodiments.

For example, a configuration in which a motive force orientation controldevice, a damping force control device, and a friction brake orientationcontrol device individually calculate control amounts that will adjustthe orientation of the vehicle body to a target orientation and controlthe engine, brakes, and variable-damping-force shock absorbers, whereinthe operation of the actuators is switched on and off according to theabsolute value of the amplitude of a detected state quantity.

1. A vehicle control device comprising: a motive power sourceorientation control device configured to calculate a motive power sourceorientation control amount for a vehicle motive power source so as toyield a target orientation for vehicle body orientation, and configuredto output the amount to the motive power source; damping force controldevice configured to calculate a shock absorber orientation controlamount for a variable-damping-force shock absorber so as to change theorientation of the vehicle body to the target orientation, andconfigured to output the shock absorber orientation control amount tothe variable-damping-force shock absorber; state quantity detectiondevice configured to detect a state quantity indicating vehicle bodyorientation; and orientation control device configured to controlvehicle body orientation via the motive power source orientation controldevice when the absolute value of the amplitude of the detected statequantity is less than a first predetermined value, via the damping forcecontrol device instead of the motive power source orientation controldevice when the absolute value of the amplitude is equal to or greaterthan the first predetermined value and less than a second predeterminedvalue greater than the first predetermined value, and via the frictionbrake orientation control device instead of the damping force controldevice when the absolute value of the amplitude is equal to or greaterthan the second predetermined value. 2-5. (canceled)
 6. The vehiclecontrol device according to claim 1, further comprising a statedetection device configured to detect vehicle pitch rate; and thefriction brake orientation control device is configured to calculate thebrake orientation control amount on the basis of the detected pitchrate.
 7. The vehicle control device according to claim 6, wherein thedriving state detection device is configured to estimate vehicle pitchrate on the basis of changes in wheel speed.
 8. The vehicle controldevice according to claim 7, wherein: the driving state detection deviceis vehicle pitch rate via a four-wheel model constructed on the basis ofa bounce term representing the four-wheel vertical movement, a pitchterm representing front and rear wheel vertical movement, a roll termrepresenting left and right wheel vertical movement, and a warp termrepresenting diagonal wheel vertical movement.
 9. The vehicle controldevice according to claim 1, wherein the orientation control device isconfigured to calculate each of the various orientation control amountson the basis of skyhook control.
 10. The vehicle control deviceaccording to claim 1, wherein the state quantity detection device isconfigured to detect vehicle pitch rate.
 11. The vehicle control deviceaccording to claim 1, wherein the state quantity detection device isconfigured to detect vehicle roll rate.
 12. A vehicle control devicecomprising: a sensor configured to detect a state quantity indicatingvehicle body orientation; and a controller configured to control vehiclebody orientation via drive force from a vehicle motive power source whenthe absolute value of the amplitude of the detected state quantity isless than a first predetermined value, via damping force from avariable-damping-force shock absorber instead of drive force from themotive power source when the absolute value of the amplitude is equal toor greater than the first predetermined value and less than a secondpredetermined value that is greater than the first predetermined value,and via braking force from a friction brake instead of damping forcefrom the variable-damping-force shock absorber when the absolute valueof the amplitude is equal to or greater than the second predeterminedvalue.
 13. A vehicle control method comprising: controlling, using acontroller, vehicle body orientation using drive force from a vehiclemotive power source when the absolute value of the amplitude of a statequantity representing vehicle body orientation is less than a firstpredetermined value, using damping force from a variable-damping-forceshock absorber instead of drive force from the motive power source whenthe absolute value of the amplitude is equal to or greater than thefirst predetermined value and less than a second predetermined valuethat is greater than the first predetermined value, and using brakingforce from a friction brake instead of damping force from thevariable-damping-force shock absorber when the absolute value of theamplitude is equal to or greater than the second predetermined value.14. The vehicle control device according to claim 8, wherein theorientation control device is configured to calculate each of theorientation control amounts on the basis of skyhook control.
 15. Thevehicle control device according to claim 9, wherein the state quantitydetection device is configured to detect vehicle pitch rate.
 16. Thevehicle control device according to claim 1, wherein the state quantitydetection device is configured to detect vehicle roll rate.